- Email: [email protected]
A thermodynamic evaluation on high pressure condenser of double effect absorption refrigeration system
Energy 113 (2016) 1031e1041
Contents lists available at ScienceDirect
Energy journal homepage: www.elsevier.com/locate/energy
A thermodynamic evaluation on high pressure condenser of double effect absorption refrigeration system _ Ibrahim Halil Yılmaz a, Kenan Saka b, Omer Kaynakli c, * a
Department of Automotive Engineering, Adana Science and Technology University, Adana, Turkey _ , Bursa, Turkey Department of Air Conditioning and Refrigeration, Vocational School of Yenis¸ehir Ibrahim Orhan, University of Uludag c , Bursa, Turkey Department of Mechanical Engineering, University of Uludag b
a r t i c l e i n f o
a b s t r a c t
Article history: Received 19 March 2016 Received in revised form 1 June 2016 Accepted 25 July 2016
One of the parameters affecting the COP of the absorption system can be considered as the thermal balance between the high pressure condenser (HPC) and the low pressure generator (LPG) since heat rejected from the HPC is utilized as an energy source by the LPG. Condensation of the water vapor in the HPC depends on the heat removal via the LPG. This circumstance is significant for making an appropriate design and a controllable system with high performance in practical applications. For this reason, a thermodynamic analysis for the HPC of a double effect series flow water/lithium bromide absorption refrigeration system was emphasized in this study. A simulation was developed to investigate the energy transfer between the HPC and LPG. The results show that the proper designation of the HPC temperature improves the COP and ECOP due its significant impact, and its value necessarily has to be higher than the outlet temperature of the LPG based on the operating scheme. Furthermore, the COP and ECOP of the absorption system can be raised in the range of 9.72e35.09% in case of 2 C-temperature increment in the HPC under the described conditions to be applied. © 2016 Elsevier Ltd. All rights reserved.
Keywords: Double effect absorption Water/lithium bromide High pressure condenser Low pressure generator Performance evaluation Irreversibility
1. Introduction In recent years, energy security and related issues have oriented us to heed the energy recovery and efficiency for thermal systems. Renewable source and waste heat utilization impress the science communities for the sake of insuring energy sustainability and curbing carbon emissions. Absorption refrigeration systems provide many distinctions from those points in cooling operations but have lower coefficient of performance (COP) relative to its counterparts. In order to improve the COP of these systems or adapt them to any source of energy, various modifications to the cycle configurations have been proposed [1e3]. Multi-effect cycles have higher COP values relative to the basic configurations however they require higher source temperature with increasing the effect number [4] and increased number of heat exchanger. On the other hand, raising the effect number is not energy effective alone when the system components are not operated in a suitable operational domain. It has been a primary challenge for the researchers to
* Corresponding author. E-mail addresses: [email protected] (H. Yılmaz), [email protected]. tr (K. Saka), [email protected] (O. Kaynakli). http://dx.doi.org/10.1016/j.energy.2016.07.133 0360-5442/© 2016 Elsevier Ltd. All rights reserved.
increase the COP of the absorption refrigeration systems to a significant degree. At this point, a theoretical analysis provides a wealth of information on the expected operational characteristic of the system and its performance. Yet, it fundamentally includes some assumptions or approximations to simplify the analysis which in turn can yield some pitfalls while not holding the physical nature of the system. Thus a realistic system analysis helps to deliver admissible outputs which serve to predict the system behavior and performance under different scenarios. Double effect absorption system, as its name implies, utilizes double-generator; high, and low pressure which in turn provide heat recovery and improvement of the COP. In double effect series flow absorption systems, it is required the entire vapor generated at the high pressure generator (HPG) to be fully condensed via the low pressure generator (LPG). This is achieved in the practical systems by installing a throttling device which allows forming the condensate in high pressure condenser (HPC) of the absorption system, and thus the noncondensate is restricted there not to be escaped to the condenser as in the vapor form [5]. At this stage, the LPG adjusts itself somehow to come to an equilibrium temperature while furnishing the complete condensation at the HPC. The evaluation of this circumstance at the design stage of the multi-effect absorption systems is crucial to sustain maximum possible heat
1032
H. Yılmaz et al. / Energy 113 (2016) 1031e1041
Nomenclature f h m_ Q_ T _ W X
circulation ratio enthalpy, kJ/kg mass flow rate, kg/s heat transfer rate, kW temperature, C power, kW solution concentration, %
Greek symbol ε effectiveness h efficiency Subscript A cr C E
absorber crystallization condenser evaporator
transfer rates at the corresponding equilibrium temperature of the HPC and other regarded system parameters. At off-design conditions, the system operates at some balanced state where optimum operating efficiency would not be obtained. Thus a realistic system analysis provides us useful knowledge to be used in the design and control of such systems during operation for attaining maximum COP at optimum condition. Many researchers have made thermodynamic analyses for improving the COP of double effect water/lithium bromide (H2O/ LiBr) absorption systems using energetic and exergetic methodologies as in the single effect [6e8]. The studies handled from the view of energetic evaluation [5,9e11] considered the effective system parameters on the COP. The analyses showed that while increasing either the HPG temperature up to a certain level or effectiveness of the solution heat exchangers (SHEs) improves the COP [5,9e11], increasing the circulation ratio [5,9,11] or solution concentration ratio [11] degrades the COP. On the other hand, many assessments have been carried out for the exergy analysis of double effect H2O/LiBr absorption refrigeration systems [12e19] since the exergetic evaluation provides more meaningful results relative to the energetic analysis. The energetic approach has a limited use owing to the fact that it is incapable of identifying the thermodynamic inefficiencies in the system components. Combining the energetic and exergetic analyses offers a better insight to the improvement of the performance and design of such systems. The recent studies have presented that the performance variation of double effect H2O/LiBr absorption refrigeration systems depends on various system parameters. Ravikumar et al. [12] showed that the ECOP (exergetic coefficient of performance) of the system increases with increasing the LPG temperature but decreases with increasing of the HPG temperature. Khaliq and Kumar [13] have indicated that increasing the HPG and evaporator temperatures leads to increasing in the COP and ECOP but decreasing in the absorber and condenser temperatures lowers both the COP and ECOP. Exergy destruction in each component of the system is also investigated in the study, and it is seen that the total exergy destruction in both the HPG and LPG make up between 30 and 35%, the SHEs correspond to 11e20%, the total exergy destruction in absorber is around 12e18% of the exergy change in the HPG heating water, and the expansion valves (EVs) are relatively smaller within the range of 1e5%. Gomri and Hakimi [14] drove similar inferences
HPC HPG LPG P s w
high pressure condenser high pressure generator low pressure generator pump; pressure strong solution weak solution
Abbreviation COP Coefficient of Performance ECOP Exergetic Coefficient of Performance EV Expansion Valve HPC High Pressure Condenser HPG High Pressure Generator H2O/LiBr Water/Lithium Bromide ITBP Initial Thermal Balance Point LPG Low Pressure Generator SHE I Solution Heat Exchanger I SHE II Solution Heat Exchanger II
with [12], and further concluded that the absorber and HPG have strong effects on system performance since their contribution to the total exergy destruction is remarkable. Additionally, the effect of the LPG temperature on the COP and ECOP is examined. Increasing the LPG temperature raises the COP and ECOP sharply but higher than a certain value does not contribute substantial improvement to the performance indicators. Kaushik and Arora [15] presented that increase in the HPG temperature increases the COP and ECOP up to an optimum HPG temperature (about 150 C). It was obtained that lowering the absorber temperature brings down the optimum HPG temperature and increases the COP and ECOP as well as their maximum values. Increasing the evaporator temperature raises the COP but reduces ECOP. Increasing the absorber temperature reduces the COP influentially as compared to increasing the condenser temperature. The absorber degrades the COP mostly in comparison to the other system components. Arora and Kaushik [16] indicated the variation of COP and ECOP with respect to the component temperatures and dead state. The results showed that the COP can be increased by employing higher HPG temperate up to a level (about 150 C) where the COP curve levels off and even shows a marginal drop at subsequent elevating temperatures. The maximum COP is obtained at lower LPG temperatures in case either the absorber/condenser temperature is reduced or the evaporator temperature and effectiveness of SHEs are increased. The optimum HPG temperature to obtain maximum ECOP was found lower (130 C for the series flow double effect system) as compared to the optimum HPG temperature that results the maximum COP. The ECOP reduces with an increase in the absorber, condenser and evaporator temperatures but the exergetic analysis introduced that the absorber is the worst component and requires more advancing design relative to the condenser, evaporator SHEs, respectively. Farshi et al. [17] compared three different cycle configurations (series, parallel and reverse parallel) with identical cooling capacities. The comparison study revealed that the COP and ECOP for the parallel and reverse parallel systems show similar trend but higher than those of the series flow. The study also resulted that an increase in HPG temperature to a certain value increased the COP and ECOP, and further increasing reduced the ECOP but remained the COP almost constant. It was also similarly found that the optimum HPG temperature corresponding to the maximum ECOP is lower than that of the corresponding
H. Yılmaz et al. / Energy 113 (2016) 1031e1041
maximum COP. Moreover, it was indicated that having a higher effectiveness in low temperature SHE improves the COP and ECOP significantly. The pressure drop between the evaporator and absorber components can be revealed as a significant design factor since its increase causes to reduce both the COP and ECOP. Kaynakli et al. [18] performed a parametric study to investigate the effect of operation temperatures of the system components on the COP and the exergy destruction of the HPG Three different heat sources (hot water, hot air and steam) were applied to the HPG, and the exergy destruction rates were compared. The maximum exergy destruction values were obtained in the hot air application, followed by the steam and hot water ones, respectively. Hamed et al. [19] predicted the performance of a double effect absorption system for evaporator temperature range of 2e10 C, absorber and condenser temperature range of 28e45 C, and HPG temperature range of 100e200 C. It was concluded that the increased temperature for the HPG increases the required heat from the HPG at a certain level after that the COP becomes almost constant and slightly falls in higher range of temperature in contrast to the ECOP. The analyses available in the literature consider the thermodynamic performance of double effect H2O/LiBr absorption refrigeration systems from the point of different aspects. Even if improving the COP and ECOP of these systems depends upon many parameters, the thermal unbalance between the HPC and LPG, and additionally the onset of the thermal balance and component heat capacities were not scrutinized for performance enhancement. The main objective of this work is to analyze double effect H2O/LiBr absorption refrigeration systems from these points. The obtained results will help to comprehend the thermal equilibrium on the HPC and its effect on the components' operating range, and the COP and ECOP change, as well. Furthermore, the presented figures will serve as a useful tool through the system design and control to raise the performance in practical applications.
1033
2. System description and methodology The schematic representation of double effect series flow H2O/ LiBr absorption refrigeration system is shown in Fig. 1. The absorption cycle uses H2O/LiBr which is a binary solution consisting of refrigerant and absorbent. The main components of the absorption system are plainly illustrated as seen. It has two steam generators known as HPG and LPG which make up the double effect. The heat supplied to the HPG is to extract the water as vapor from the solution. The vapor is condensed in the HPC, and the water in the diluted solution is vaporized again by the LPG. The SHE I and II are used herein to increase the performance of the absorption system. Unlike the vapor-compression systems, the absorption system uses an absorber to absorb the vaporized refrigerant by the concentrated LiBr solution. The thermodynamic analysis of the absorption system described is made theoretically in the previous study [18] applying the principles of mass, energy and exergy balance. It is noted that the analysis includes the following simplifications and approximations which are appropriate from the view of practical system operation.
Steady-state operating conditions exist. The evaporator has a fixed capacity. Pressure losses in heat exchangers and pipelines are negligible. Water phase is saturated liquid and saturated vapor at the outlet of the condenser and evaporator, respectively. Heat exchange between the system components and surroundings is disregarded. Only heat input/output is taken into account for the prescribed system components. The dead state temperature and pressure are assumed to be at 25 C and 1 atm, respectively. Calculation of ECOP is estimated by the equation proposed by Ref. [16]. The temperature of the cold source in the evaporator
Fig. 1. Double effect series flow absorption refrigeration system.
1034
H. Yılmaz et al. / Energy 113 (2016) 1031e1041
and the hot source in the HPG are taken at the averaged temperature.
3. Solution approximation and model validation The concerned equations mentioned in Section 2 involve many parameters. The prediction of the thermodynamic properties of the working fluids (H2O, and H2O/LiBr solution) is performed by the correlations provided by Refs. [20e22]. To solve all the model
equations, estimation of the thermal balance temperature in the HPC/LPG unit is required. The approaches for describing this temperature were considered in various ways in the literature including T12 ¼ T15 [10,15,16,24], DT1215 ¼ fixed [5,13] and T12 s T15 [9,11,14,17,18]. The more realistic approach is the case of T12 s T15 [23] which was used as a solution approach in this study since it presents a better physical insight into the system. For this object, a computer program was written in Delphi to predict the unknowns by the solution algorithm given in Fig. 2. Validation of the simulation study was fulfilled by comparing
Fig. 2. Systematic flowchart for modeling of double effect absorption system [25].
H. Yılmaz et al. / Energy 113 (2016) 1031e1041 Table 1 Comparison of the simulation results to the reference work. Component
Arora and Kaushik [16]
Present study
Difference (%)
HPG, kW LPG, kW Condenser, kW Absorber, kW Evaporator, kW Pump, kW SHE I, kW SHE II, kW COP
1868.71 1272.48 1282.05 2942.18 2355.45 0.3598 518.59 816.21 1.26
1863.20 1276.88 1262.45 2922.58 2355.45 0.4382 524.11 842.60 1.26
0.29 0.34 1.53 0.67 e 21.79 1.06 3.23 e
Parameter
HPG
THPG ( C) Outlet/Inlet TLPG ( C) TC ( C) Outlet/Inlet TA ( C) Outlet/Inlet TE ( C) Outlet/Inlet Q_ (kW)
LPG Condenser Absorber Evaporator
Operating value temperature of hot water
temperature of cooling water temperature of cooling water temperature of chilled water
E
Pump SHE I, SHE II
hP (%) εI,II (%)
130e140 THPG þ 5/THPG þ 15 75e85 33e39 TC5/TC10 33e39 TA5/TA10 4e12 TE þ 5/TE þ 10 100 95 70
the results to the reference study of [16]. The following operating parameters were employed to the developed program as THPG ¼ 138.15 C, TE ¼ 7.2 C, TC ¼ TA ¼ 37.8 C, εI,II ¼ 0.7 and m_ w ¼ 1:0kg=s. The outputs of the simulation are given in Table 1 providing the case of T12 ¼ T15 since this approach was used in TA = 36 C, TC = 36 C, TE = 8 C,
I,II
= 0.7
TLPG = 80 C, TC = 36 C, TE = 8 C,
115
TLPG= 75 C TLPG= 80 C
105
In this section, the effects of the variation in the HPC temperature on component capacity, the onset of the thermal balance and the path of quasi-equilibrium, COP and ECOP change, and exergy destruction are introduced in terms of varying component temperatures. The reference operating parameters and their range tabulated in Table 2 were applied to the system otherwise the opposite is not noted. The figures presented through the section involve a set of variable parameters. Each satisfies the absorption system to work within the confident operational domain, i.e., considering solution concentration, crystallization risk and freezing. The inlet and outlet temperatures of hot, cooling and chilled waters were considered as a function of the temperature of the main components to maintain solely the energy balance on the related component. Fig. 3 shows the variation in the HPC temperature as a function of HPG temperature. In each figure, the effect of varying temperature of the LPG, condenser, absorber and evaporator is illustrated respectively. It is seen that there is a linear relationship between the HPG and HPC temperatures. In Fig. 3a, the growing LPG temperature allows the absorption system to operate at the higher HPG and HPC temperatures. This gives rise to more vapor generation in the HPG and increases the heat capacity of the LPG as a consequence. The onset of the thermal balance initiates from a
HPC temperature ( C)
HPC temperature ( C)
115
that study. As it is seen, the present model yields good agreement as compared to the reference work. Here, the difference for the solution pump has been obtained bigger relative to the reference work. It is related that the solution pump possesses a small capacity hence a little change in its value produces a big deviation. On the other hand, many studies in the _ P z0 literature disregard the pump work [5,10e12] or consider W [14,24]. 4. Results and discussion
Table 2 Operating parameters used in the simulation. Component
TLPG= 85 C
95 85 75
I,II
= 0.7
TA= 33 C TA= 36 C
105
TA= 39 C
95 85 75
115
125
135 145 HPG temperature ( C)
155
165
115
125
(a) TLPG = 80 C, TA = 36 C, TE = 8 C,
115
TLPG = 80 C, TA = 36 C, TC = 36 C,
115
TC= 33 C
TC= 36 C
105
135 145 HPG temperature ( C)
155
165
(b) = 0.7 I,II
HPC temperature ( C)
HPC temperature ( C)
1035
TC= 39 C
95 85
I,II
= 0.7
TE= 4 C TE= 8 C
105
TE= 12 C
95 85
75
75 115
125
135 145 HPG temperature ( C)
(c)
155
165
115
125
135 145 HPG temperature ( C)
(d)
Fig. 3. Variation in HPC temperature based on HPG and component temperatures.
155
165
1036
H. Yılmaz et al. / Energy 113 (2016) 1031e1041 I,II =
0.7
THPG= 135 C, TLPG = 80 C, TC = 36 C, TE = 8 C,
3000
2500
2500
2000 1500 QHPC QLPG
1000
THPG= 130 C THPG= 135 C THPG= 140 C
500 0 78
80
82
Heat capacity (kJ/kg)
Heat capacity (kJ/kg)
TLPG = 80 C, TA = 36 C, TC = 36 C, TE = 8 C,
3000
92
1500 QHPC QLPG
1000
TA= 33 C TA= 36 C TA= 39 C
500
94
78
80
82
(a) 3000
2500
2500
2000
1500 QHPC QLPG TC= 33 C TC= 36 C TC= 39 C
0 78
80
82
92
THPG= 135 C, TLPG = 80 C, TC = 36 C, TA = 36 C,
Heat capacity (kJ/kg)
Heat capacity (kJ/kg)
I,II = 0.7
3000
500
84 86 88 90 HPC temperature ( C)
94
(b)
THPG= 135 C, TLPG = 80 C, TA = 36 C, TE = 8 C,
1000
0.7
2000
0
84 86 88 90 HPC temperature ( C)
I,II =
92
0.7
2000 1500 QHPC QLPG
1000
TE= 4 C TE= 8 C TE= 12 C
500 0
84 86 88 90 HPC temperature ( C)
I,II =
94
78
80
(c)
82
84 86 88 90 HPC temperature ( C)
92
94
(d)
Fig. 4. Variation in heat capacities of LPG and HPC based on HPC and component temperatures.
TLPG = 80 C, TA = 36 C, TC = 36 C, TE = 8 C,
I,II
= 0.7
THPG= 130 C THPG= 135 C
4300
THPG = 135 C, TLPG= 80 C, TC = 36 C, TE = 8 C,
4800 HPG capacity (kJ/kg)
HPG capacity (kJ/kg)
4800
THPG= 140 C
3800 3300 2800
= 0.7
TA= 33 C TA= 36 C
4300
TA= 39 C
3800 3300 2800
80
82
84 86 88 90 HPC temperature ( C)
92
94
80
82
(a) = 0.7 I,II
92
TC= 36 C
TC= 39 C
3800
3300 2800
THPG = 135 C, TLPG= 80 C, TA = 36 C, TC = 36 C,
4800 HPG capacity (kJ/kg)
TC= 33 C
4300
84 86 88 90 HPC temperature ( C)
94
(b)
THPG = 135 C, TLPG= 80 C, TA = 36 C, TE = 8 C,
4800 HPG capacity (kJ/kg)
I,II
I,II
= 0.7
TE= 4 C TE= 8 C
4300
TE= 12 C
3800 3300 2800
80
82
84 86 88 90 HPC temperature ( C)
92
94
(c)
80
82
84 86 88 90 HPC temperature ( C)
92
94
(d)
Fig. 5. Variation in heat capacity of HPG based on HPC and component temperatures.
minimum temperature (closely higher than the LPG temperature) and then reaches a maximum value. Increasing the absorber
temperature, as it is seen from Fig. 3b, raises the initial thermal balance point (ITBP) of the HPC. Contrarily, increasing of both the
H. Yılmaz et al. / Energy 113 (2016) 1031e1041
condenser and evaporator temperatures lowers this point as shown in Fig. 3c and d. It is inferred that the minimum HPC temperature satisfying the thermal balance must be always higher than the LPG temperature. Fig. 4 exhibits the tracing path of quasi-equilibrium for the heat capacities of the LPG and HPC, and its variation depending on the thermal balance temperature of the HPC also is shown. It is noted that all heat capacities given throughout this study are expressed in kJ/kg since two circulation ratios used in the analysis to define the heat capacities of the components without using the flow rate of the working fluid. The heat capacity rate can be calculated multiplying the heat capacity with the vapor generated by the HPG, i.e., m_ 11 . It is seen that increasing the HPC temperature raises the capacity of the LPG significantly but declines the capacity of the HPC slightly on the condition of elevating both the HPG and LPG temperatures (see Fig. 4aed). The slight variation in the HPC capacity originates from the enthalpy difference between the HPG and HPC outlets since the enthalpy falls at state 11 and increases at state 12 for sustaining the condition of thermal balance. Obtaining maximum possible heat capacity from the LPG requires higher HPC temperature operation. Increasing the absorber temperature limits the maximum HPC temperature whereas the effects of raised condenser and evaporator temperatures cause the absorption system to operate at higher HPC temperature. Further increasing the condenser temperature shifts the ITBP temperature while the LPG temperature is kept constant. The variation in the HPG capacity in terms of HPC and component temperature is given in Fig. 5. Increasing the HPG shifts the HPC temperature to a higher degree which results in slight capacity growth in the HPG. It is seen that the HPG capacity does not strongly depend on the HPG temperature in case of elevating it under the conditions applied. While the ITBP is 80.01 C at THPG@130 C, it climbs to 87.94 C for the case of THPG@140 C although the LPG temperature is set to 80 C as seen in Fig. 5a. Increasing the
Absorber capacity (kJ/kg)
= 0.7
I,II
4700 THPG= 130 C THPG= 135 C THPG= 140 C
2700 80
82
84 86 88 90 HPC temperature ( C)
THPG = 135 C, TLPG = 80 C, TC = 36 C, TE = 8 C,
6700
5700
3700
absorber temperature raises the HPG capacity remarkably. The ITBP moves away from the LPG temperature for each temperature increment in the absorber, and the absorber temperature begins to fall the HPG capacity as seen in Fig. 5b. Increasing the condenser temperature extends the HPG capacity to be upward sloping with rise of the HPC temperature as seen in Fig. 5c. At TC@33 C, the HPG capacity changes slowly but higher the condensing temperature engenders more rapid climbing in the HPG capacity. In order to decrease the HPG capacity, the condenser and HPC should be operated relatively at low temperatures since there is a direct relation changing polynomially. Fig. 5d shows that increasing the evaporator temperature gets flatten the capacity but affects it inversely as compared to the case of Fig. 5b, and widens the operational range of the HPC device. In Fig. 6, the variation in the absorber capacity in terms of HPC and component temperature is illustrated. It is shown in Fig. 6a that increasing the HPG temperature does not affect the capacity of the absorber. On the other hand, raising the absorber temperature increases the absorber capacity as a function of the HPC temperature shown in Fig. 6b. At the highest absorber temperature, the absorber capacity increases dramatically. 3 C-temperature increment in the absorber almost doubles the absorber capacity. The curves obtained are upward sloping, and their slope increases with increasing of the HPC temperature. The absorber capacity can be increased more in case of raising the condenser temperature as seen from Fig. 6c. The evaporator capacity change is inversely affected by the increased evaporator temperature since there is an opposite relation between the absorber and evaporator capacities as demonstrated in Fig. 6b and d. The variation of the condenser capacity versus HPC and component temperature is given through Fig. 7aed. The increase of both the HPG and LPG temperatures reduces the condenser capacity as clearly seen from Fig. 7a, however it increases with increasing the absorber temperature. The ITBP for each case is very
Absorber capacity (kJ/kg)
TLPG = 80 C, TA = 36 C, TC = 36 C, TE = 8 C,
6700
92
TA= 33 C
3700
TA= 36 C TA= 39 C
80
94
82
TC= 39 C
82
84 86 88 90 HPC temperature ( C)
(c)
92
92
THPG = 135 C, TLPG = 80 C, TC= 36 C, TE = 8 C,
6700
Absorber capacity (kJ/kg)
Absorber capacity (kJ/kg)
TC= 33 C TC= 36 C
80
84 86 88 90 HPC temperature ( C)
94
(b) = 0.7 I,II
4700
2700
= 0.7
4700
2700
5700
3700
I,II
5700
(a) THPG = 135 C, TLPG = 80 C, TA= 36 C, TE = 8 C,
6700
1037
= 0.7
5700 4700 TE= 4 C
3700
TE= 8 C TE= 12 C
2700 94
I,II
80
82
84 86 88 90 HPC temperature ( C)
(d)
Fig. 6. Variation in heat capacity of absorber based on HPC and component temperatures.
92
94
H. Yılmaz et al. / Energy 113 (2016) 1031e1041
TLPG = 80 C, TA = 36 C, TC = 36 C, TE = 8 C,
Condenser capacity (kJ/kg)
3000
I,II
= 0.7
2500 2000 1500
1000
THPG= 130 C
500
THPG= 135 C
THPG= 140 C
0 80
82
84 86 88 90 HPC temperature ( C)
THPG = 135 C, TLPG = 80 C, TC = 36 C, TE = 8 C,
3000
Condenser capacity (kJ/kg)
1038
92
2000 1500 1000
TA= 33 C
500
TA= 36 C TA= 39 C
0
94
80
82
1500 1000
TC= 33 C
500
TC= 36 C
TC= 39 C
82
92
84 86 88 90 HPC temperature ( C)
92
THPG = 135 C, TLPG = 80 C, TC= 36 C, TE = 8 C,
3000
Condenser capacity (kJ/kg)
Condenser capacity (kJ/kg)
2000
80
84 86 88 90 HPC temperature ( C)
94
(b) = 0.7 I,II
2500
0
= 0.7
2500
(a) THPG = 135 C, TLPG = 80 C, TA = 36 C, TE = 8 C,
3000
I,II
I,II
= 0.7
2500 2000 1500
1000
TE= 4 C
500
TE= 8 C TE= 12 C
0 80
94
82
(c)
84 86 88 90 HPC temperature ( C)
92
94
(d)
Fig. 7. Variation in heat capacity of condenser based on HPC and component temperatures.
Evaporator capacity (kJ/kg)
I,II
= 0.7
4500 4000 3500
3000
THPG= 130 C
2500
THPG= 135 C
THPG= 140 C
2000 80
82
84 86 88 90 HPC temperature ( C)
THPG = 135 C, TLPG = 80 C, TC = 36 C, TE = 8 C,
5000
Evaporator capacity (kJ/kg)
TLPG = 80 C, TA = 36 C, TC = 36 C, TE = 8 C,
5000
92
4000 3500 3000
TA= 33 C TA= 36 C
2500
TA= 39 C
2000
94
80
82
3500 3000
TC= 33 C
2500
TC= 36 C
TC= 39 C
82
92
84 86 88 90 HPC temperature ( C)
92
94
(c)
THPG = 135 C, TLPG = 80 C, TC= 36 C, TE = 8 C,
5000
Evaporator capacity (kJ/kg)
Evaporator capacity (kJ/kg)
4000
80
84 86 88 90 HPC temperature ( C)
94
(b) = 0.7 I,II
4500
2000
= 0.7
4500
(a) THPG = 135 C, TLPG = 80 C, TA = 36 C, TE = 8 C,
5000
I,II
I,II
= 0.7
4500
4000 3500 3000
TE= 4 C TE= 8 C
2500
TE= 12 C
2000 80
82
84 86 88 90 HPC temperature ( C)
92
94
(d)
Fig. 8. Variation in heat capacity of evaporator based on HPC and component temperatures.
close to each other as pointed in Fig. 7b and d. The condenser capacity is seriously affected by the increased HPC temperature. It grows more than 6 times once the highest HPC temperature is
attained. This illustrates the sensitivity of the condenser capacity to the change of the HPC temperature. The variation of the evaporator capacity depending on HPC and
H. Yılmaz et al. / Energy 113 (2016) 1031e1041
TLPG = 80 C, TA = 36 C, TC = 36 C, TE = 8 C,
I,II
= 0.7
THPG = 135 C, TLPG = 80 C, TC = 36 C, TE = 8 C,
1.4
1.2
1.2
1
1
COP
COP
1.4
0.8
1039
0.8 TA= 33 C
THPG= 130 C
0.6
0.6
THPG= 135 C THPG= 140 C
0.4 80
82
84 86 88 90 HPC temperature ( C)
92
TA= 36 C TA= 39 C
0.4 94
80
82
(a) = 0.7 I,II
1.2
1.2
1
1
0.8
0.6
TC= 39 C
0.4 84 86 88 90 HPC temperature ( C)
92
I,II
= 0.7
TE= 4 C
0.6
TC= 36 C
82
94
0.8
TC= 33 C
80
92
THPG = 135 C, TLPG = 80 C, TA = 36 C, TC = 36 C,
1.4
COP
COP
84 86 88 90 HPC temperature ( C)
(b)
THPG = 135 C, TLPG = 80 C, TA = 36 C, TE = 8 C,
1.4
= 0.7
I,II
TE= 8 C
TE= 12 C
0.4 80
94
82
(c)
84 86 88 90 HPC temperature ( C)
92
94
(d)
Fig. 9. Variation of COP based on HPG and component temperatures.
TLPG = 80 C, TA = 36 C, TC = 36 C, TE = 8 C,
20
I,II
= 0.7
10 THPG= 130 C
5
= 0.7
THPG= 140 C
80
82
84 86 88 90 HPC temperature ( C)
92
10 TA= 33 C
5
THPG= 135 C
0
TA= 36 C TA= 39 C
0 94
80
82
(a)
84 86 88 90 HPC temperature ( C)
92
94
(b)
THPG = 135 C, TLPG = 80 C, TA = 36 C, TE = 8 C,
20
= 0.7 I,II
THPG = 135 C, TLPG = 80 C, TA = 36 C, TC = 36 C,
20
I,II
= 0.7
15
ECOP (%)
15
ECOP (%)
I,II
15
ECOP (%)
15
ECOP (%)
THPG = 135 C, TLPG = 80 C, TC = 36 C, TE = 8 C,
20
10 TC= 33 C
5
TC= 39 C
80
82
84 86 88 90 HPC temperature ( C)
92
TE= 4 C
5
TC= 36 C
0
10
TE= 8 C TE= 12 C
0 94
(c)
80
82
84 86 88 90 HPC temperature ( C)
92
94
(d)
Fig. 10. Variation of ECOP based on HPG and component temperatures.
component temperature is shown through Fig. 8aed. Obtained figures show similar trends with Figs. 6 and 7, only the capacities differ from each other. The evaporator capacity almost doubles with
increasing of the HPC temperature. The increase in the capacity change for the evaporator and absorber is nearly the same. Not only changing the component temperatures but also the thermal
1040
H. Yılmaz et al. / Energy 113 (2016) 1031e1041
balance within the HPC/LPG unit is an effective method to increase the performance. Thus the thermal balance at the HPC has an impact on the components capacities and accordingly the resulting performance. The COP variation corresponding to the HPC and component temperatures are shown in Fig. 9. As it is expected increasing only the HPG decreases the COP of the absorption system but its effect is seen weak. To obtain higher COP from the system, it is needed to rise up the HPC temperature as deduced from Fig. 9a for the given set of parameters. The increased absorber temperature has a negative effect on the COP since it narrows the operational limits of the HPC. However, the COP rise at those higher absorber temperatures is more rapid relative to the low ones. The maximum COP is obtained 1.330 with [email protected] C under the given conditions as seen in Fig. 9b. The COP variation at TA@39 C is 0.46 as the HPC temperature rises from [email protected] C to [email protected] C. This change is equal to 0.55 at TA@33 C once the HPC temperature rises from [email protected] C to [email protected] C. 3 C-temperature increment in the absorber temperature increases the HPC temperature range as much as the absorber’. As it is known, increasing the condenser temperature lowers the COP since it does not change the maximum condenser capacity remarkably. Furthermore, increasing the condenser temperature causes to grow the HPG capacity whose increase degrades the COP as shown in Fig. 9c. However, it should be noted that increasing condenser temperature necessitates higher HPC temperatures. As it is inferred from Fig. 9a, lowering the HPG enhances the COP but requires higher HPC temperatures. The maximum COP is obtained 1.324 at [email protected] C and TC@33 C for the given conditions. The effect of the evaporator temperature is shown in Fig. 9d. The maximum COP value is obtained 1.372 at TE@12 C with [email protected] C. It is expected that raising the evaporator capacity with lesser HPG input boosts the COP. This is possible at high evaporator and HPC temperatures as seen. Furthermore, the effect of the evaporator temperature on the COP is the most decisive parameter as compared with the other component temperatures. Its increase offers a wide operation gap to the system operation. The COP can be raised minimally 9.72%, and 35.09% maximally in case of 2 C-temperature increment in the HPC under the conditions of TE@12 C and TE@4 C, respectively. This increase for the temperature change of the HPG, absorber, and condenser is in the interval of 12.77e19.41%, 9.87e23.30%, and 10.60e23.62%, respectively. Although increasing the absorber temperature has a negative effect on the COP, increasing the evaporator temperature assists positively it; the COP value is surprisingly higher at high absorber temperature relative to high evaporator temperature level for the same HPC temperature. Thus, the obtained figures present to clarify the importance of the thermal balance temperature inside the HPC/LPG unit from the aspect of resulting component capacities and COP. The variation of ECOP with HPC temperature at varying component temperatures is shown in Fig. 10. The ECOP increases with an increase in the HPC temperature for all conditions (see Fig. 10aed). Increasing THPG from 130 C to 140 C decreases the maximum COP change only 2.20% but the ECOP lowers 7.86%. The reason why the ECOP falls much more relative to the COP is that increased HPG temperature influences the ECOP negatively as clearly inferred from the ECOP relation. When the HPG temperature is experienced 10 C-difference, the minimum and maximum ECOP are in the range of 8.46e15.25% as deduced from Fig. 10a. This range for 6 C-difference in the absorber and condenser temperatures are in the interval of 8.50e15.24% and 8.48e15.17%, respectively as seen from the related figures. It is understood that 10 C-difference in the HPG corresponds to almost same ECOP values with 6 C-difference in the absorber and condenser temperatures. However, that range is in the level of 5.18e16.89% for 6 C-difference in the evaporator temperature. It can be seen that the system results in
Table 3 Heat capacities of system components.a Component
Capacity (kW)
HPG LPG Condenser Absorber Evaporator Pump SHE I SHE II
78.22 51.42 55.82 122.53 100.00 0.019 23.07 42.40
a THPG ¼ 135 C, THPC ¼ 89.5 C, TLPG ¼ 80 C, TA ¼ 36 C, TC ¼ 36 C, TE ¼ 8 C, εI,II ¼ 0.7.
Fig. 11. Exergy destruction of system components.
better ECOP at low HPG, absorber, condenser and evaporator temperatures within increased HPC temperatures. The ECOP can be raised at the same rate with the COP for the case of 2 C-temperature increment in the HPC under the same conditions. It is also indicated that the ECOP is highly dependent on the evaporator temperature rather than HPG temperature. The heat loads of the system components are given in Table 3 under the operating parameters. The maximum capacity rate has yielded by the absorber while the solution pump has the lowest. These data are given for analyzing the exergy destruction of the components easily. The exergy destructions for each component of the absorption system are shown in Fig. 11 to compare the irreversibilities. The total exergy destruction under the given conditions is 19.01 kW of which 34.43% is the highest fraction donated by the absorber, and followed by the evaporator and HPG among the components. The main reason is, in general, that the absorber involves a highly inherent irreversible process i.e., mixing of the water vapor and solution [26]. The exergy destruction in the pump is the smallest one to be neglected. This order is related to the capacity largeness of the components and the operational parameters. However the component capacities are taken into consideration, 5.34% of the absorber's capacity has lost due to irreversibility, and it is followed by the SHE II, SHE I, HPG, condenser, evaporator, and LPG to be fractionally as 4.65%, 3.84%, 3.69%, 3.14%, 3.11% and 2.40%, respectively. 5. Conclusions The heat exchange between the HPC and LPG components is a significant event. The outlet temperature of the HPC cannot be lower thermodynamically than the outlet temperature of the LPG. However, the thermal balance between these components affects
H. Yılmaz et al. / Energy 113 (2016) 1031e1041
the performance of the absorption system. In this work, the heat transfer balance between the HPC and LPG and the corresponding HPC temperatures satisfying the system operation were studied since the COP and ECOP are heavily affected by those corresponding parameters as well. A simulation program has been developed in Delphi to make a parametric analysis of the absorption system. The obtained results indicate that the ITBP for the HPC has to be constantly higher than the outlet temperature of the LPG for maximizing the capacity of the LPG and accordingly the COP and ECOP. Furthermore, the small temperature changes in the system components can be effective on the performance indicators of the system as explicitly shown thus the system control and operational management are absolutely significant concerns in practical applications. References [1] Srikhirin P, Aphornratana S, Chungpaibulpatana S. A review of absorption refrigeration technologies. Renew Sustain Energy Rev 2001;5:343e72. [2] Xu ZY, Wang RZ, Xia ZZ. A novel variable effect LiBr-water absorption refrigeration cycle. Energy 2013;60:457e63. € Investigation on double effect dual-heated _ [3] Saka K, Yılmaz IH, Kaynaklı O. absorption refrigeration system. XII. International HVACþR Technology _ Symposium, Istanbul, 123e129, MareApr 2016. [4] Berhane HG, Medrano M, Boer D. Exergy analysis of multi-effect watereLiBr absorption systems: from half to triple effect. Renew Energy 2010;35: 1773e82. [5] Arun MB, Maiya MP, Murthy SS. Equilibrium low pressure generator temperatures for double-effect series flow absorption refrigeration systems. Appl Therm Eng 2000;20:227e42. [6] Kilic M, Kaynakli O. Second law-based thermodynamic analysis of waterlithium bromide absorption refrigeration system. Energy 2007;32:1505e12. [7] Wonchala J, Hazledine M, Boulama KG. Solution procedure and performance evaluation for a watereLiBr absorption refrigeration machine. Energy 2014;65:272e84. [8] Gong S, Boulama KG. Parametric study of an absorption refrigeration machine using advanced exergy analysis. Energy 2014;76:453e67. [9] Vliet GC, Lawson MB, Lithgow RA. Waterelithium bromide double effect absorption cooling system analysis. ASHRAE Trans 1982;5:811e23.
1041
[10] Kaushik SC, Chandra S. Computer modeling and parametric study of a double effect generation absorption refrigeration cycle. Energy Convers Manag 1982;25:9e14. [11] Xu GP, Dai YQ, Tou KW, Tso CP. Theoretical analysis and optimization of a double-effect series-flow-type absorption chiller. Appl Therm Eng 1996;16: 975e87. [12] Ravikumar TS, Suganthi L, Samuel AA. Exergy analysis of solar assisted double effect absorption refrigeration system. Renew Energy 1998;14:55e9. [13] Khaliq A, Kumar R. Exergy analysis of double effect vapor absorption refrigeration system. Int J Energy Res 2008;32:161e74. [14] Gomri R, Hakimi R. Second law analysis of double effect vapour absorption cooler system. Energy Convers Manag 2008;49:3343e8. [15] Kaushik SC, Arora A. Energy and exergy analysis of single effect and series flow double effect waterelithium bromide absorption refrigeration systems. Int J Refrig 2009;32:1247e58. [16] Arora A, Kaushik SC. Theoretical analysis of LiBr/H2O absorption refrigeration systems. Int J Energy Res 2009;33:1321e40. [17] Farshi LG, Mahmoudi SMS, Rosen MA, Yari M. A comparative study of the performance characteristics of double-effect absorption refrigeration systems. Int J Energy Res 2012;36:182e92. [18] Kaynakli O, Saka K, Kaynakli F. Energy and exergy analysis of a double effect absorption refrigeration system based on different heat sources. Energy Convers Manag 2015;106:21e30. [19] Hamed A, Kaseb SA, Hanafi AS. Prediction of energetic and exergetic performance of double-effect absorption system. Int J Hydrogen Energy 2015;40: 15320e7. [20] Kaita Y. Thermodynamic properties of lithium bromide-water solutions at high temperatures. Int J Refrig 2001;24:374e90. [21] Mostafavi M, Agnew B. The impact of ambient temperature on lithium bromideewater absorption machine performance. Appl Therm Eng 1996;16: 515e22. [22] Chua HT, Toh HK, Malek A, Ng KC, Srinivasan K. Improved thermodynamic property field of LiBreH2O solution. Int J Refrig 2000;23:412e29. € Influence of the equilibrium temperature in the _ Saka K, Kaynaklı O. [23] Yılmaz IH, double stage of the absorption refrigeration system. 8th International Ege Energy Symposium and Exhibition, Afyon, May 2016. [24] Anand DK, Kumar B. Absorption machine irreversibility using new entropy calculations. Sol Energy 1987;39:243e56. [25] Kaynakli O, Kilic M. Theoretical study on the effect of operating conditions on performance of absorption refrigeration system. Energy Convers Manag 2007;48:599e607. [26] Kaynakli O. The first and second law analysis of a lithium bromide/water coil absorber. Energy 2008;33:804e16.
Contents lists available at ScienceDirect
Energy journal homepage: www.elsevier.com/locate/energy
A thermodynamic evaluation on high pressure condenser of double effect absorption refrigeration system _ Ibrahim Halil Yılmaz a, Kenan Saka b, Omer Kaynakli c, * a
Department of Automotive Engineering, Adana Science and Technology University, Adana, Turkey _ , Bursa, Turkey Department of Air Conditioning and Refrigeration, Vocational School of Yenis¸ehir Ibrahim Orhan, University of Uludag c , Bursa, Turkey Department of Mechanical Engineering, University of Uludag b
a r t i c l e i n f o
a b s t r a c t
Article history: Received 19 March 2016 Received in revised form 1 June 2016 Accepted 25 July 2016
One of the parameters affecting the COP of the absorption system can be considered as the thermal balance between the high pressure condenser (HPC) and the low pressure generator (LPG) since heat rejected from the HPC is utilized as an energy source by the LPG. Condensation of the water vapor in the HPC depends on the heat removal via the LPG. This circumstance is significant for making an appropriate design and a controllable system with high performance in practical applications. For this reason, a thermodynamic analysis for the HPC of a double effect series flow water/lithium bromide absorption refrigeration system was emphasized in this study. A simulation was developed to investigate the energy transfer between the HPC and LPG. The results show that the proper designation of the HPC temperature improves the COP and ECOP due its significant impact, and its value necessarily has to be higher than the outlet temperature of the LPG based on the operating scheme. Furthermore, the COP and ECOP of the absorption system can be raised in the range of 9.72e35.09% in case of 2 C-temperature increment in the HPC under the described conditions to be applied. © 2016 Elsevier Ltd. All rights reserved.
Keywords: Double effect absorption Water/lithium bromide High pressure condenser Low pressure generator Performance evaluation Irreversibility
1. Introduction In recent years, energy security and related issues have oriented us to heed the energy recovery and efficiency for thermal systems. Renewable source and waste heat utilization impress the science communities for the sake of insuring energy sustainability and curbing carbon emissions. Absorption refrigeration systems provide many distinctions from those points in cooling operations but have lower coefficient of performance (COP) relative to its counterparts. In order to improve the COP of these systems or adapt them to any source of energy, various modifications to the cycle configurations have been proposed [1e3]. Multi-effect cycles have higher COP values relative to the basic configurations however they require higher source temperature with increasing the effect number [4] and increased number of heat exchanger. On the other hand, raising the effect number is not energy effective alone when the system components are not operated in a suitable operational domain. It has been a primary challenge for the researchers to
* Corresponding author. E-mail addresses: [email protected] (H. Yılmaz), [email protected]. tr (K. Saka), [email protected] (O. Kaynakli). http://dx.doi.org/10.1016/j.energy.2016.07.133 0360-5442/© 2016 Elsevier Ltd. All rights reserved.
increase the COP of the absorption refrigeration systems to a significant degree. At this point, a theoretical analysis provides a wealth of information on the expected operational characteristic of the system and its performance. Yet, it fundamentally includes some assumptions or approximations to simplify the analysis which in turn can yield some pitfalls while not holding the physical nature of the system. Thus a realistic system analysis helps to deliver admissible outputs which serve to predict the system behavior and performance under different scenarios. Double effect absorption system, as its name implies, utilizes double-generator; high, and low pressure which in turn provide heat recovery and improvement of the COP. In double effect series flow absorption systems, it is required the entire vapor generated at the high pressure generator (HPG) to be fully condensed via the low pressure generator (LPG). This is achieved in the practical systems by installing a throttling device which allows forming the condensate in high pressure condenser (HPC) of the absorption system, and thus the noncondensate is restricted there not to be escaped to the condenser as in the vapor form [5]. At this stage, the LPG adjusts itself somehow to come to an equilibrium temperature while furnishing the complete condensation at the HPC. The evaluation of this circumstance at the design stage of the multi-effect absorption systems is crucial to sustain maximum possible heat
1032
H. Yılmaz et al. / Energy 113 (2016) 1031e1041
Nomenclature f h m_ Q_ T _ W X
circulation ratio enthalpy, kJ/kg mass flow rate, kg/s heat transfer rate, kW temperature, C power, kW solution concentration, %
Greek symbol ε effectiveness h efficiency Subscript A cr C E
absorber crystallization condenser evaporator
transfer rates at the corresponding equilibrium temperature of the HPC and other regarded system parameters. At off-design conditions, the system operates at some balanced state where optimum operating efficiency would not be obtained. Thus a realistic system analysis provides us useful knowledge to be used in the design and control of such systems during operation for attaining maximum COP at optimum condition. Many researchers have made thermodynamic analyses for improving the COP of double effect water/lithium bromide (H2O/ LiBr) absorption systems using energetic and exergetic methodologies as in the single effect [6e8]. The studies handled from the view of energetic evaluation [5,9e11] considered the effective system parameters on the COP. The analyses showed that while increasing either the HPG temperature up to a certain level or effectiveness of the solution heat exchangers (SHEs) improves the COP [5,9e11], increasing the circulation ratio [5,9,11] or solution concentration ratio [11] degrades the COP. On the other hand, many assessments have been carried out for the exergy analysis of double effect H2O/LiBr absorption refrigeration systems [12e19] since the exergetic evaluation provides more meaningful results relative to the energetic analysis. The energetic approach has a limited use owing to the fact that it is incapable of identifying the thermodynamic inefficiencies in the system components. Combining the energetic and exergetic analyses offers a better insight to the improvement of the performance and design of such systems. The recent studies have presented that the performance variation of double effect H2O/LiBr absorption refrigeration systems depends on various system parameters. Ravikumar et al. [12] showed that the ECOP (exergetic coefficient of performance) of the system increases with increasing the LPG temperature but decreases with increasing of the HPG temperature. Khaliq and Kumar [13] have indicated that increasing the HPG and evaporator temperatures leads to increasing in the COP and ECOP but decreasing in the absorber and condenser temperatures lowers both the COP and ECOP. Exergy destruction in each component of the system is also investigated in the study, and it is seen that the total exergy destruction in both the HPG and LPG make up between 30 and 35%, the SHEs correspond to 11e20%, the total exergy destruction in absorber is around 12e18% of the exergy change in the HPG heating water, and the expansion valves (EVs) are relatively smaller within the range of 1e5%. Gomri and Hakimi [14] drove similar inferences
HPC HPG LPG P s w
high pressure condenser high pressure generator low pressure generator pump; pressure strong solution weak solution
Abbreviation COP Coefficient of Performance ECOP Exergetic Coefficient of Performance EV Expansion Valve HPC High Pressure Condenser HPG High Pressure Generator H2O/LiBr Water/Lithium Bromide ITBP Initial Thermal Balance Point LPG Low Pressure Generator SHE I Solution Heat Exchanger I SHE II Solution Heat Exchanger II
with [12], and further concluded that the absorber and HPG have strong effects on system performance since their contribution to the total exergy destruction is remarkable. Additionally, the effect of the LPG temperature on the COP and ECOP is examined. Increasing the LPG temperature raises the COP and ECOP sharply but higher than a certain value does not contribute substantial improvement to the performance indicators. Kaushik and Arora [15] presented that increase in the HPG temperature increases the COP and ECOP up to an optimum HPG temperature (about 150 C). It was obtained that lowering the absorber temperature brings down the optimum HPG temperature and increases the COP and ECOP as well as their maximum values. Increasing the evaporator temperature raises the COP but reduces ECOP. Increasing the absorber temperature reduces the COP influentially as compared to increasing the condenser temperature. The absorber degrades the COP mostly in comparison to the other system components. Arora and Kaushik [16] indicated the variation of COP and ECOP with respect to the component temperatures and dead state. The results showed that the COP can be increased by employing higher HPG temperate up to a level (about 150 C) where the COP curve levels off and even shows a marginal drop at subsequent elevating temperatures. The maximum COP is obtained at lower LPG temperatures in case either the absorber/condenser temperature is reduced or the evaporator temperature and effectiveness of SHEs are increased. The optimum HPG temperature to obtain maximum ECOP was found lower (130 C for the series flow double effect system) as compared to the optimum HPG temperature that results the maximum COP. The ECOP reduces with an increase in the absorber, condenser and evaporator temperatures but the exergetic analysis introduced that the absorber is the worst component and requires more advancing design relative to the condenser, evaporator SHEs, respectively. Farshi et al. [17] compared three different cycle configurations (series, parallel and reverse parallel) with identical cooling capacities. The comparison study revealed that the COP and ECOP for the parallel and reverse parallel systems show similar trend but higher than those of the series flow. The study also resulted that an increase in HPG temperature to a certain value increased the COP and ECOP, and further increasing reduced the ECOP but remained the COP almost constant. It was also similarly found that the optimum HPG temperature corresponding to the maximum ECOP is lower than that of the corresponding
H. Yılmaz et al. / Energy 113 (2016) 1031e1041
maximum COP. Moreover, it was indicated that having a higher effectiveness in low temperature SHE improves the COP and ECOP significantly. The pressure drop between the evaporator and absorber components can be revealed as a significant design factor since its increase causes to reduce both the COP and ECOP. Kaynakli et al. [18] performed a parametric study to investigate the effect of operation temperatures of the system components on the COP and the exergy destruction of the HPG Three different heat sources (hot water, hot air and steam) were applied to the HPG, and the exergy destruction rates were compared. The maximum exergy destruction values were obtained in the hot air application, followed by the steam and hot water ones, respectively. Hamed et al. [19] predicted the performance of a double effect absorption system for evaporator temperature range of 2e10 C, absorber and condenser temperature range of 28e45 C, and HPG temperature range of 100e200 C. It was concluded that the increased temperature for the HPG increases the required heat from the HPG at a certain level after that the COP becomes almost constant and slightly falls in higher range of temperature in contrast to the ECOP. The analyses available in the literature consider the thermodynamic performance of double effect H2O/LiBr absorption refrigeration systems from the point of different aspects. Even if improving the COP and ECOP of these systems depends upon many parameters, the thermal unbalance between the HPC and LPG, and additionally the onset of the thermal balance and component heat capacities were not scrutinized for performance enhancement. The main objective of this work is to analyze double effect H2O/LiBr absorption refrigeration systems from these points. The obtained results will help to comprehend the thermal equilibrium on the HPC and its effect on the components' operating range, and the COP and ECOP change, as well. Furthermore, the presented figures will serve as a useful tool through the system design and control to raise the performance in practical applications.
1033
2. System description and methodology The schematic representation of double effect series flow H2O/ LiBr absorption refrigeration system is shown in Fig. 1. The absorption cycle uses H2O/LiBr which is a binary solution consisting of refrigerant and absorbent. The main components of the absorption system are plainly illustrated as seen. It has two steam generators known as HPG and LPG which make up the double effect. The heat supplied to the HPG is to extract the water as vapor from the solution. The vapor is condensed in the HPC, and the water in the diluted solution is vaporized again by the LPG. The SHE I and II are used herein to increase the performance of the absorption system. Unlike the vapor-compression systems, the absorption system uses an absorber to absorb the vaporized refrigerant by the concentrated LiBr solution. The thermodynamic analysis of the absorption system described is made theoretically in the previous study [18] applying the principles of mass, energy and exergy balance. It is noted that the analysis includes the following simplifications and approximations which are appropriate from the view of practical system operation.
Steady-state operating conditions exist. The evaporator has a fixed capacity. Pressure losses in heat exchangers and pipelines are negligible. Water phase is saturated liquid and saturated vapor at the outlet of the condenser and evaporator, respectively. Heat exchange between the system components and surroundings is disregarded. Only heat input/output is taken into account for the prescribed system components. The dead state temperature and pressure are assumed to be at 25 C and 1 atm, respectively. Calculation of ECOP is estimated by the equation proposed by Ref. [16]. The temperature of the cold source in the evaporator
Fig. 1. Double effect series flow absorption refrigeration system.
1034
H. Yılmaz et al. / Energy 113 (2016) 1031e1041
and the hot source in the HPG are taken at the averaged temperature.
3. Solution approximation and model validation The concerned equations mentioned in Section 2 involve many parameters. The prediction of the thermodynamic properties of the working fluids (H2O, and H2O/LiBr solution) is performed by the correlations provided by Refs. [20e22]. To solve all the model
equations, estimation of the thermal balance temperature in the HPC/LPG unit is required. The approaches for describing this temperature were considered in various ways in the literature including T12 ¼ T15 [10,15,16,24], DT1215 ¼ fixed [5,13] and T12 s T15 [9,11,14,17,18]. The more realistic approach is the case of T12 s T15 [23] which was used as a solution approach in this study since it presents a better physical insight into the system. For this object, a computer program was written in Delphi to predict the unknowns by the solution algorithm given in Fig. 2. Validation of the simulation study was fulfilled by comparing
Fig. 2. Systematic flowchart for modeling of double effect absorption system [25].
H. Yılmaz et al. / Energy 113 (2016) 1031e1041 Table 1 Comparison of the simulation results to the reference work. Component
Arora and Kaushik [16]
Present study
Difference (%)
HPG, kW LPG, kW Condenser, kW Absorber, kW Evaporator, kW Pump, kW SHE I, kW SHE II, kW COP
1868.71 1272.48 1282.05 2942.18 2355.45 0.3598 518.59 816.21 1.26
1863.20 1276.88 1262.45 2922.58 2355.45 0.4382 524.11 842.60 1.26
0.29 0.34 1.53 0.67 e 21.79 1.06 3.23 e
Parameter
HPG
THPG ( C) Outlet/Inlet TLPG ( C) TC ( C) Outlet/Inlet TA ( C) Outlet/Inlet TE ( C) Outlet/Inlet Q_ (kW)
LPG Condenser Absorber Evaporator
Operating value temperature of hot water
temperature of cooling water temperature of cooling water temperature of chilled water
E
Pump SHE I, SHE II
hP (%) εI,II (%)
130e140 THPG þ 5/THPG þ 15 75e85 33e39 TC5/TC10 33e39 TA5/TA10 4e12 TE þ 5/TE þ 10 100 95 70
the results to the reference study of [16]. The following operating parameters were employed to the developed program as THPG ¼ 138.15 C, TE ¼ 7.2 C, TC ¼ TA ¼ 37.8 C, εI,II ¼ 0.7 and m_ w ¼ 1:0kg=s. The outputs of the simulation are given in Table 1 providing the case of T12 ¼ T15 since this approach was used in TA = 36 C, TC = 36 C, TE = 8 C,
I,II
= 0.7
TLPG = 80 C, TC = 36 C, TE = 8 C,
115
TLPG= 75 C TLPG= 80 C
105
In this section, the effects of the variation in the HPC temperature on component capacity, the onset of the thermal balance and the path of quasi-equilibrium, COP and ECOP change, and exergy destruction are introduced in terms of varying component temperatures. The reference operating parameters and their range tabulated in Table 2 were applied to the system otherwise the opposite is not noted. The figures presented through the section involve a set of variable parameters. Each satisfies the absorption system to work within the confident operational domain, i.e., considering solution concentration, crystallization risk and freezing. The inlet and outlet temperatures of hot, cooling and chilled waters were considered as a function of the temperature of the main components to maintain solely the energy balance on the related component. Fig. 3 shows the variation in the HPC temperature as a function of HPG temperature. In each figure, the effect of varying temperature of the LPG, condenser, absorber and evaporator is illustrated respectively. It is seen that there is a linear relationship between the HPG and HPC temperatures. In Fig. 3a, the growing LPG temperature allows the absorption system to operate at the higher HPG and HPC temperatures. This gives rise to more vapor generation in the HPG and increases the heat capacity of the LPG as a consequence. The onset of the thermal balance initiates from a
HPC temperature ( C)
HPC temperature ( C)
115
that study. As it is seen, the present model yields good agreement as compared to the reference work. Here, the difference for the solution pump has been obtained bigger relative to the reference work. It is related that the solution pump possesses a small capacity hence a little change in its value produces a big deviation. On the other hand, many studies in the _ P z0 literature disregard the pump work [5,10e12] or consider W [14,24]. 4. Results and discussion
Table 2 Operating parameters used in the simulation. Component
TLPG= 85 C
95 85 75
I,II
= 0.7
TA= 33 C TA= 36 C
105
TA= 39 C
95 85 75
115
125
135 145 HPG temperature ( C)
155
165
115
125
(a) TLPG = 80 C, TA = 36 C, TE = 8 C,
115
TLPG = 80 C, TA = 36 C, TC = 36 C,
115
TC= 33 C
TC= 36 C
105
135 145 HPG temperature ( C)
155
165
(b) = 0.7 I,II
HPC temperature ( C)
HPC temperature ( C)
1035
TC= 39 C
95 85
I,II
= 0.7
TE= 4 C TE= 8 C
105
TE= 12 C
95 85
75
75 115
125
135 145 HPG temperature ( C)
(c)
155
165
115
125
135 145 HPG temperature ( C)
(d)
Fig. 3. Variation in HPC temperature based on HPG and component temperatures.
155
165
1036
H. Yılmaz et al. / Energy 113 (2016) 1031e1041 I,II =
0.7
THPG= 135 C, TLPG = 80 C, TC = 36 C, TE = 8 C,
3000
2500
2500
2000 1500 QHPC QLPG
1000
THPG= 130 C THPG= 135 C THPG= 140 C
500 0 78
80
82
Heat capacity (kJ/kg)
Heat capacity (kJ/kg)
TLPG = 80 C, TA = 36 C, TC = 36 C, TE = 8 C,
3000
92
1500 QHPC QLPG
1000
TA= 33 C TA= 36 C TA= 39 C
500
94
78
80
82
(a) 3000
2500
2500
2000
1500 QHPC QLPG TC= 33 C TC= 36 C TC= 39 C
0 78
80
82
92
THPG= 135 C, TLPG = 80 C, TC = 36 C, TA = 36 C,
Heat capacity (kJ/kg)
Heat capacity (kJ/kg)
I,II = 0.7
3000
500
84 86 88 90 HPC temperature ( C)
94
(b)
THPG= 135 C, TLPG = 80 C, TA = 36 C, TE = 8 C,
1000
0.7
2000
0
84 86 88 90 HPC temperature ( C)
I,II =
92
0.7
2000 1500 QHPC QLPG
1000
TE= 4 C TE= 8 C TE= 12 C
500 0
84 86 88 90 HPC temperature ( C)
I,II =
94
78
80
(c)
82
84 86 88 90 HPC temperature ( C)
92
94
(d)
Fig. 4. Variation in heat capacities of LPG and HPC based on HPC and component temperatures.
TLPG = 80 C, TA = 36 C, TC = 36 C, TE = 8 C,
I,II
= 0.7
THPG= 130 C THPG= 135 C
4300
THPG = 135 C, TLPG= 80 C, TC = 36 C, TE = 8 C,
4800 HPG capacity (kJ/kg)
HPG capacity (kJ/kg)
4800
THPG= 140 C
3800 3300 2800
= 0.7
TA= 33 C TA= 36 C
4300
TA= 39 C
3800 3300 2800
80
82
84 86 88 90 HPC temperature ( C)
92
94
80
82
(a) = 0.7 I,II
92
TC= 36 C
TC= 39 C
3800
3300 2800
THPG = 135 C, TLPG= 80 C, TA = 36 C, TC = 36 C,
4800 HPG capacity (kJ/kg)
TC= 33 C
4300
84 86 88 90 HPC temperature ( C)
94
(b)
THPG = 135 C, TLPG= 80 C, TA = 36 C, TE = 8 C,
4800 HPG capacity (kJ/kg)
I,II
I,II
= 0.7
TE= 4 C TE= 8 C
4300
TE= 12 C
3800 3300 2800
80
82
84 86 88 90 HPC temperature ( C)
92
94
(c)
80
82
84 86 88 90 HPC temperature ( C)
92
94
(d)
Fig. 5. Variation in heat capacity of HPG based on HPC and component temperatures.
minimum temperature (closely higher than the LPG temperature) and then reaches a maximum value. Increasing the absorber
temperature, as it is seen from Fig. 3b, raises the initial thermal balance point (ITBP) of the HPC. Contrarily, increasing of both the
H. Yılmaz et al. / Energy 113 (2016) 1031e1041
condenser and evaporator temperatures lowers this point as shown in Fig. 3c and d. It is inferred that the minimum HPC temperature satisfying the thermal balance must be always higher than the LPG temperature. Fig. 4 exhibits the tracing path of quasi-equilibrium for the heat capacities of the LPG and HPC, and its variation depending on the thermal balance temperature of the HPC also is shown. It is noted that all heat capacities given throughout this study are expressed in kJ/kg since two circulation ratios used in the analysis to define the heat capacities of the components without using the flow rate of the working fluid. The heat capacity rate can be calculated multiplying the heat capacity with the vapor generated by the HPG, i.e., m_ 11 . It is seen that increasing the HPC temperature raises the capacity of the LPG significantly but declines the capacity of the HPC slightly on the condition of elevating both the HPG and LPG temperatures (see Fig. 4aed). The slight variation in the HPC capacity originates from the enthalpy difference between the HPG and HPC outlets since the enthalpy falls at state 11 and increases at state 12 for sustaining the condition of thermal balance. Obtaining maximum possible heat capacity from the LPG requires higher HPC temperature operation. Increasing the absorber temperature limits the maximum HPC temperature whereas the effects of raised condenser and evaporator temperatures cause the absorption system to operate at higher HPC temperature. Further increasing the condenser temperature shifts the ITBP temperature while the LPG temperature is kept constant. The variation in the HPG capacity in terms of HPC and component temperature is given in Fig. 5. Increasing the HPG shifts the HPC temperature to a higher degree which results in slight capacity growth in the HPG. It is seen that the HPG capacity does not strongly depend on the HPG temperature in case of elevating it under the conditions applied. While the ITBP is 80.01 C at THPG@130 C, it climbs to 87.94 C for the case of THPG@140 C although the LPG temperature is set to 80 C as seen in Fig. 5a. Increasing the
Absorber capacity (kJ/kg)
= 0.7
I,II
4700 THPG= 130 C THPG= 135 C THPG= 140 C
2700 80
82
84 86 88 90 HPC temperature ( C)
THPG = 135 C, TLPG = 80 C, TC = 36 C, TE = 8 C,
6700
5700
3700
absorber temperature raises the HPG capacity remarkably. The ITBP moves away from the LPG temperature for each temperature increment in the absorber, and the absorber temperature begins to fall the HPG capacity as seen in Fig. 5b. Increasing the condenser temperature extends the HPG capacity to be upward sloping with rise of the HPC temperature as seen in Fig. 5c. At TC@33 C, the HPG capacity changes slowly but higher the condensing temperature engenders more rapid climbing in the HPG capacity. In order to decrease the HPG capacity, the condenser and HPC should be operated relatively at low temperatures since there is a direct relation changing polynomially. Fig. 5d shows that increasing the evaporator temperature gets flatten the capacity but affects it inversely as compared to the case of Fig. 5b, and widens the operational range of the HPC device. In Fig. 6, the variation in the absorber capacity in terms of HPC and component temperature is illustrated. It is shown in Fig. 6a that increasing the HPG temperature does not affect the capacity of the absorber. On the other hand, raising the absorber temperature increases the absorber capacity as a function of the HPC temperature shown in Fig. 6b. At the highest absorber temperature, the absorber capacity increases dramatically. 3 C-temperature increment in the absorber almost doubles the absorber capacity. The curves obtained are upward sloping, and their slope increases with increasing of the HPC temperature. The absorber capacity can be increased more in case of raising the condenser temperature as seen from Fig. 6c. The evaporator capacity change is inversely affected by the increased evaporator temperature since there is an opposite relation between the absorber and evaporator capacities as demonstrated in Fig. 6b and d. The variation of the condenser capacity versus HPC and component temperature is given through Fig. 7aed. The increase of both the HPG and LPG temperatures reduces the condenser capacity as clearly seen from Fig. 7a, however it increases with increasing the absorber temperature. The ITBP for each case is very
Absorber capacity (kJ/kg)
TLPG = 80 C, TA = 36 C, TC = 36 C, TE = 8 C,
6700
92
TA= 33 C
3700
TA= 36 C TA= 39 C
80
94
82
TC= 39 C
82
84 86 88 90 HPC temperature ( C)
(c)
92
92
THPG = 135 C, TLPG = 80 C, TC= 36 C, TE = 8 C,
6700
Absorber capacity (kJ/kg)
Absorber capacity (kJ/kg)
TC= 33 C TC= 36 C
80
84 86 88 90 HPC temperature ( C)
94
(b) = 0.7 I,II
4700
2700
= 0.7
4700
2700
5700
3700
I,II
5700
(a) THPG = 135 C, TLPG = 80 C, TA= 36 C, TE = 8 C,
6700
1037
= 0.7
5700 4700 TE= 4 C
3700
TE= 8 C TE= 12 C
2700 94
I,II
80
82
84 86 88 90 HPC temperature ( C)
(d)
Fig. 6. Variation in heat capacity of absorber based on HPC and component temperatures.
92
94
H. Yılmaz et al. / Energy 113 (2016) 1031e1041
TLPG = 80 C, TA = 36 C, TC = 36 C, TE = 8 C,
Condenser capacity (kJ/kg)
3000
I,II
= 0.7
2500 2000 1500
1000
THPG= 130 C
500
THPG= 135 C
THPG= 140 C
0 80
82
84 86 88 90 HPC temperature ( C)
THPG = 135 C, TLPG = 80 C, TC = 36 C, TE = 8 C,
3000
Condenser capacity (kJ/kg)
1038
92
2000 1500 1000
TA= 33 C
500
TA= 36 C TA= 39 C
0
94
80
82
1500 1000
TC= 33 C
500
TC= 36 C
TC= 39 C
82
92
84 86 88 90 HPC temperature ( C)
92
THPG = 135 C, TLPG = 80 C, TC= 36 C, TE = 8 C,
3000
Condenser capacity (kJ/kg)
Condenser capacity (kJ/kg)
2000
80
84 86 88 90 HPC temperature ( C)
94
(b) = 0.7 I,II
2500
0
= 0.7
2500
(a) THPG = 135 C, TLPG = 80 C, TA = 36 C, TE = 8 C,
3000
I,II
I,II
= 0.7
2500 2000 1500
1000
TE= 4 C
500
TE= 8 C TE= 12 C
0 80
94
82
(c)
84 86 88 90 HPC temperature ( C)
92
94
(d)
Fig. 7. Variation in heat capacity of condenser based on HPC and component temperatures.
Evaporator capacity (kJ/kg)
I,II
= 0.7
4500 4000 3500
3000
THPG= 130 C
2500
THPG= 135 C
THPG= 140 C
2000 80
82
84 86 88 90 HPC temperature ( C)
THPG = 135 C, TLPG = 80 C, TC = 36 C, TE = 8 C,
5000
Evaporator capacity (kJ/kg)
TLPG = 80 C, TA = 36 C, TC = 36 C, TE = 8 C,
5000
92
4000 3500 3000
TA= 33 C TA= 36 C
2500
TA= 39 C
2000
94
80
82
3500 3000
TC= 33 C
2500
TC= 36 C
TC= 39 C
82
92
84 86 88 90 HPC temperature ( C)
92
94
(c)
THPG = 135 C, TLPG = 80 C, TC= 36 C, TE = 8 C,
5000
Evaporator capacity (kJ/kg)
Evaporator capacity (kJ/kg)
4000
80
84 86 88 90 HPC temperature ( C)
94
(b) = 0.7 I,II
4500
2000
= 0.7
4500
(a) THPG = 135 C, TLPG = 80 C, TA = 36 C, TE = 8 C,
5000
I,II
I,II
= 0.7
4500
4000 3500 3000
TE= 4 C TE= 8 C
2500
TE= 12 C
2000 80
82
84 86 88 90 HPC temperature ( C)
92
94
(d)
Fig. 8. Variation in heat capacity of evaporator based on HPC and component temperatures.
close to each other as pointed in Fig. 7b and d. The condenser capacity is seriously affected by the increased HPC temperature. It grows more than 6 times once the highest HPC temperature is
attained. This illustrates the sensitivity of the condenser capacity to the change of the HPC temperature. The variation of the evaporator capacity depending on HPC and
H. Yılmaz et al. / Energy 113 (2016) 1031e1041
TLPG = 80 C, TA = 36 C, TC = 36 C, TE = 8 C,
I,II
= 0.7
THPG = 135 C, TLPG = 80 C, TC = 36 C, TE = 8 C,
1.4
1.2
1.2
1
1
COP
COP
1.4
0.8
1039
0.8 TA= 33 C
THPG= 130 C
0.6
0.6
THPG= 135 C THPG= 140 C
0.4 80
82
84 86 88 90 HPC temperature ( C)
92
TA= 36 C TA= 39 C
0.4 94
80
82
(a) = 0.7 I,II
1.2
1.2
1
1
0.8
0.6
TC= 39 C
0.4 84 86 88 90 HPC temperature ( C)
92
I,II
= 0.7
TE= 4 C
0.6
TC= 36 C
82
94
0.8
TC= 33 C
80
92
THPG = 135 C, TLPG = 80 C, TA = 36 C, TC = 36 C,
1.4
COP
COP
84 86 88 90 HPC temperature ( C)
(b)
THPG = 135 C, TLPG = 80 C, TA = 36 C, TE = 8 C,
1.4
= 0.7
I,II
TE= 8 C
TE= 12 C
0.4 80
94
82
(c)
84 86 88 90 HPC temperature ( C)
92
94
(d)
Fig. 9. Variation of COP based on HPG and component temperatures.
TLPG = 80 C, TA = 36 C, TC = 36 C, TE = 8 C,
20
I,II
= 0.7
10 THPG= 130 C
5
= 0.7
THPG= 140 C
80
82
84 86 88 90 HPC temperature ( C)
92
10 TA= 33 C
5
THPG= 135 C
0
TA= 36 C TA= 39 C
0 94
80
82
(a)
84 86 88 90 HPC temperature ( C)
92
94
(b)
THPG = 135 C, TLPG = 80 C, TA = 36 C, TE = 8 C,
20
= 0.7 I,II
THPG = 135 C, TLPG = 80 C, TA = 36 C, TC = 36 C,
20
I,II
= 0.7
15
ECOP (%)
15
ECOP (%)
I,II
15
ECOP (%)
15
ECOP (%)
THPG = 135 C, TLPG = 80 C, TC = 36 C, TE = 8 C,
20
10 TC= 33 C
5
TC= 39 C
80
82
84 86 88 90 HPC temperature ( C)
92
TE= 4 C
5
TC= 36 C
0
10
TE= 8 C TE= 12 C
0 94
(c)
80
82
84 86 88 90 HPC temperature ( C)
92
94
(d)
Fig. 10. Variation of ECOP based on HPG and component temperatures.
component temperature is shown through Fig. 8aed. Obtained figures show similar trends with Figs. 6 and 7, only the capacities differ from each other. The evaporator capacity almost doubles with
increasing of the HPC temperature. The increase in the capacity change for the evaporator and absorber is nearly the same. Not only changing the component temperatures but also the thermal
1040
H. Yılmaz et al. / Energy 113 (2016) 1031e1041
balance within the HPC/LPG unit is an effective method to increase the performance. Thus the thermal balance at the HPC has an impact on the components capacities and accordingly the resulting performance. The COP variation corresponding to the HPC and component temperatures are shown in Fig. 9. As it is expected increasing only the HPG decreases the COP of the absorption system but its effect is seen weak. To obtain higher COP from the system, it is needed to rise up the HPC temperature as deduced from Fig. 9a for the given set of parameters. The increased absorber temperature has a negative effect on the COP since it narrows the operational limits of the HPC. However, the COP rise at those higher absorber temperatures is more rapid relative to the low ones. The maximum COP is obtained 1.330 with [email protected] C under the given conditions as seen in Fig. 9b. The COP variation at TA@39 C is 0.46 as the HPC temperature rises from [email protected] C to [email protected] C. This change is equal to 0.55 at TA@33 C once the HPC temperature rises from [email protected] C to [email protected] C. 3 C-temperature increment in the absorber temperature increases the HPC temperature range as much as the absorber’. As it is known, increasing the condenser temperature lowers the COP since it does not change the maximum condenser capacity remarkably. Furthermore, increasing the condenser temperature causes to grow the HPG capacity whose increase degrades the COP as shown in Fig. 9c. However, it should be noted that increasing condenser temperature necessitates higher HPC temperatures. As it is inferred from Fig. 9a, lowering the HPG enhances the COP but requires higher HPC temperatures. The maximum COP is obtained 1.324 at [email protected] C and TC@33 C for the given conditions. The effect of the evaporator temperature is shown in Fig. 9d. The maximum COP value is obtained 1.372 at TE@12 C with [email protected] C. It is expected that raising the evaporator capacity with lesser HPG input boosts the COP. This is possible at high evaporator and HPC temperatures as seen. Furthermore, the effect of the evaporator temperature on the COP is the most decisive parameter as compared with the other component temperatures. Its increase offers a wide operation gap to the system operation. The COP can be raised minimally 9.72%, and 35.09% maximally in case of 2 C-temperature increment in the HPC under the conditions of TE@12 C and TE@4 C, respectively. This increase for the temperature change of the HPG, absorber, and condenser is in the interval of 12.77e19.41%, 9.87e23.30%, and 10.60e23.62%, respectively. Although increasing the absorber temperature has a negative effect on the COP, increasing the evaporator temperature assists positively it; the COP value is surprisingly higher at high absorber temperature relative to high evaporator temperature level for the same HPC temperature. Thus, the obtained figures present to clarify the importance of the thermal balance temperature inside the HPC/LPG unit from the aspect of resulting component capacities and COP. The variation of ECOP with HPC temperature at varying component temperatures is shown in Fig. 10. The ECOP increases with an increase in the HPC temperature for all conditions (see Fig. 10aed). Increasing THPG from 130 C to 140 C decreases the maximum COP change only 2.20% but the ECOP lowers 7.86%. The reason why the ECOP falls much more relative to the COP is that increased HPG temperature influences the ECOP negatively as clearly inferred from the ECOP relation. When the HPG temperature is experienced 10 C-difference, the minimum and maximum ECOP are in the range of 8.46e15.25% as deduced from Fig. 10a. This range for 6 C-difference in the absorber and condenser temperatures are in the interval of 8.50e15.24% and 8.48e15.17%, respectively as seen from the related figures. It is understood that 10 C-difference in the HPG corresponds to almost same ECOP values with 6 C-difference in the absorber and condenser temperatures. However, that range is in the level of 5.18e16.89% for 6 C-difference in the evaporator temperature. It can be seen that the system results in
Table 3 Heat capacities of system components.a Component
Capacity (kW)
HPG LPG Condenser Absorber Evaporator Pump SHE I SHE II
78.22 51.42 55.82 122.53 100.00 0.019 23.07 42.40
a THPG ¼ 135 C, THPC ¼ 89.5 C, TLPG ¼ 80 C, TA ¼ 36 C, TC ¼ 36 C, TE ¼ 8 C, εI,II ¼ 0.7.
Fig. 11. Exergy destruction of system components.
better ECOP at low HPG, absorber, condenser and evaporator temperatures within increased HPC temperatures. The ECOP can be raised at the same rate with the COP for the case of 2 C-temperature increment in the HPC under the same conditions. It is also indicated that the ECOP is highly dependent on the evaporator temperature rather than HPG temperature. The heat loads of the system components are given in Table 3 under the operating parameters. The maximum capacity rate has yielded by the absorber while the solution pump has the lowest. These data are given for analyzing the exergy destruction of the components easily. The exergy destructions for each component of the absorption system are shown in Fig. 11 to compare the irreversibilities. The total exergy destruction under the given conditions is 19.01 kW of which 34.43% is the highest fraction donated by the absorber, and followed by the evaporator and HPG among the components. The main reason is, in general, that the absorber involves a highly inherent irreversible process i.e., mixing of the water vapor and solution [26]. The exergy destruction in the pump is the smallest one to be neglected. This order is related to the capacity largeness of the components and the operational parameters. However the component capacities are taken into consideration, 5.34% of the absorber's capacity has lost due to irreversibility, and it is followed by the SHE II, SHE I, HPG, condenser, evaporator, and LPG to be fractionally as 4.65%, 3.84%, 3.69%, 3.14%, 3.11% and 2.40%, respectively. 5. Conclusions The heat exchange between the HPC and LPG components is a significant event. The outlet temperature of the HPC cannot be lower thermodynamically than the outlet temperature of the LPG. However, the thermal balance between these components affects
H. Yılmaz et al. / Energy 113 (2016) 1031e1041
the performance of the absorption system. In this work, the heat transfer balance between the HPC and LPG and the corresponding HPC temperatures satisfying the system operation were studied since the COP and ECOP are heavily affected by those corresponding parameters as well. A simulation program has been developed in Delphi to make a parametric analysis of the absorption system. The obtained results indicate that the ITBP for the HPC has to be constantly higher than the outlet temperature of the LPG for maximizing the capacity of the LPG and accordingly the COP and ECOP. Furthermore, the small temperature changes in the system components can be effective on the performance indicators of the system as explicitly shown thus the system control and operational management are absolutely significant concerns in practical applications. References [1] Srikhirin P, Aphornratana S, Chungpaibulpatana S. A review of absorption refrigeration technologies. Renew Sustain Energy Rev 2001;5:343e72. [2] Xu ZY, Wang RZ, Xia ZZ. A novel variable effect LiBr-water absorption refrigeration cycle. Energy 2013;60:457e63. € Investigation on double effect dual-heated _ [3] Saka K, Yılmaz IH, Kaynaklı O. absorption refrigeration system. XII. International HVACþR Technology _ Symposium, Istanbul, 123e129, MareApr 2016. [4] Berhane HG, Medrano M, Boer D. Exergy analysis of multi-effect watereLiBr absorption systems: from half to triple effect. Renew Energy 2010;35: 1773e82. [5] Arun MB, Maiya MP, Murthy SS. Equilibrium low pressure generator temperatures for double-effect series flow absorption refrigeration systems. Appl Therm Eng 2000;20:227e42. [6] Kilic M, Kaynakli O. Second law-based thermodynamic analysis of waterlithium bromide absorption refrigeration system. Energy 2007;32:1505e12. [7] Wonchala J, Hazledine M, Boulama KG. Solution procedure and performance evaluation for a watereLiBr absorption refrigeration machine. Energy 2014;65:272e84. [8] Gong S, Boulama KG. Parametric study of an absorption refrigeration machine using advanced exergy analysis. Energy 2014;76:453e67. [9] Vliet GC, Lawson MB, Lithgow RA. Waterelithium bromide double effect absorption cooling system analysis. ASHRAE Trans 1982;5:811e23.
1041
[10] Kaushik SC, Chandra S. Computer modeling and parametric study of a double effect generation absorption refrigeration cycle. Energy Convers Manag 1982;25:9e14. [11] Xu GP, Dai YQ, Tou KW, Tso CP. Theoretical analysis and optimization of a double-effect series-flow-type absorption chiller. Appl Therm Eng 1996;16: 975e87. [12] Ravikumar TS, Suganthi L, Samuel AA. Exergy analysis of solar assisted double effect absorption refrigeration system. Renew Energy 1998;14:55e9. [13] Khaliq A, Kumar R. Exergy analysis of double effect vapor absorption refrigeration system. Int J Energy Res 2008;32:161e74. [14] Gomri R, Hakimi R. Second law analysis of double effect vapour absorption cooler system. Energy Convers Manag 2008;49:3343e8. [15] Kaushik SC, Arora A. Energy and exergy analysis of single effect and series flow double effect waterelithium bromide absorption refrigeration systems. Int J Refrig 2009;32:1247e58. [16] Arora A, Kaushik SC. Theoretical analysis of LiBr/H2O absorption refrigeration systems. Int J Energy Res 2009;33:1321e40. [17] Farshi LG, Mahmoudi SMS, Rosen MA, Yari M. A comparative study of the performance characteristics of double-effect absorption refrigeration systems. Int J Energy Res 2012;36:182e92. [18] Kaynakli O, Saka K, Kaynakli F. Energy and exergy analysis of a double effect absorption refrigeration system based on different heat sources. Energy Convers Manag 2015;106:21e30. [19] Hamed A, Kaseb SA, Hanafi AS. Prediction of energetic and exergetic performance of double-effect absorption system. Int J Hydrogen Energy 2015;40: 15320e7. [20] Kaita Y. Thermodynamic properties of lithium bromide-water solutions at high temperatures. Int J Refrig 2001;24:374e90. [21] Mostafavi M, Agnew B. The impact of ambient temperature on lithium bromideewater absorption machine performance. Appl Therm Eng 1996;16: 515e22. [22] Chua HT, Toh HK, Malek A, Ng KC, Srinivasan K. Improved thermodynamic property field of LiBreH2O solution. Int J Refrig 2000;23:412e29. € Influence of the equilibrium temperature in the _ Saka K, Kaynaklı O. [23] Yılmaz IH, double stage of the absorption refrigeration system. 8th International Ege Energy Symposium and Exhibition, Afyon, May 2016. [24] Anand DK, Kumar B. Absorption machine irreversibility using new entropy calculations. Sol Energy 1987;39:243e56. [25] Kaynakli O, Kilic M. Theoretical study on the effect of operating conditions on performance of absorption refrigeration system. Energy Convers Manag 2007;48:599e607. [26] Kaynakli O. The first and second law analysis of a lithium bromide/water coil absorber. Energy 2008;33:804e16.